1. Field of the Invention
The present invention generally relates to a valve timing control mechanism for an internal combustion engine, which controls the timing to open or close intake and/or exhaust valves of the internal combustion engine in response to the operating condition of the engine.
2. Related Art
As is well known in an internal combustion engine, crankshaft revolution is transmitted by an appropriate transmission mechanism to a cam shaft on which a cam is disposed in order to open and close an intake valve and/or exhaust valve of the engine. It is also well known that the these crankshaft and cam shaft are usually maintained at a relative angular velocity ratio of 2:1 by means of the transmission mechanism.
Here, the open or close timing of the intake valve and exhaust valve, which are set depending on the crankshaft, transmission mechanism and cam shaft, are considered to be set at an optimum timing at the rated output, for example, of the internal combustion engine. However, if the open or close timing is set fixedly like that, the timing tends to displace minutely under a low or high load operation of the engine so that such an inconvenience as a back flow of exhaust gas may occur.
FIG. 7 shows an example of a conventional valve timing control mechanism which uses a rotational phase adjusting mechanism as explained below. The mechanism of FIG. 7 is applied to a DOHC (Double Overhead Cam) type internal combustion engine and constructed as a mechanism to variably control the open or close timing of intake valves of the engine.
First of all, the DOHC type internal combustion engine 10 is so arranged that the revolution of crankshaft 11 is transmitted to each cam shaft 15 and 16 for the exhaust valve and intake valve via the transmission mechanism consisting of each sprocket for the exhaust valve and intake valve and a timing chain 14. As described previously, that these crankshaft 11 and cam shafts 15 and 16 are maintained at the rotation angular velocity of 2:1 by means of the transmission mechanism.
Further, on the mechanism of FIG. 7, a rotational phase adjusting mechanism 20 (shadowed section in the figure) is disposed between the intake valve sprocket 13 and the cam shaft 16 and the intake valve open or close timing control is executed, in accordance with an actuating amount provided by the adjusting mechanism 20 in the following manner:
(A) The cam shaft 16 is angularly advanced relative to the sprocket 13, i.e., the crankshaft 11, PA1 (B) Relative rotation angle of the sprocket 13, i.e., the crankshaft 11 and the cam shaft 16 is retained, and PA1 (C) The cam shaft 16 is angularly delayed relative to the sprocket 13, i.e., the crankshaft 11.
Here, a hydraulic mechanism 50 basically comprises an oil pan 51 which reserves working oil, an oil pump 52 which pumps the reserved working oil, and a spool valve 53 which distributes the pumped working oil to each hydraulic chamber, which is referred to later, of the rotational phase adjusting mechanism 20.
In addition, in the hydraulic mechanism 50, the working oil distribution movement by the spool valve 53 is arranged to be controlled via a linear solenoid 56. Drive of the linear solenoid 56 is controlled by a duty ratio of electric current applied by an electronic control unit 70 and the control of angular advance of the cam shaft 16 in step (A) above is realized by the duty value of the current in a direction to increase it from 50% while, on the contrary, the control of angular delay of the cam shaft 16 in step (C) is realized in a direction to reduce it from 50%.
Further, on the mechanism of FIG. 7, a crank angle sensor 17, which detects the rotational angle of the crankshaft 11, is installed near the crankshaft 11 of the internal combustion engine 10 and, similarly, a cam angle sensor 18 to detect the rotational angle of the cam shaft 16 is installed near the cam shaft 16.
Both of these angle sensors 17 and 18 are well known sensors which comprise of a plural number of bodies (lugs) subject to detection which are made of, for example, a magnetic material and an electromagnetic pickup sensor to detect the rotating passage of these detected bodies, and each of these transmits pulse signals corresponding to the rotational angle of each shaft to the control unit 70.
In this connection, the crank angle sensor 17 of the same mechanism generates four pulse signals per revolution of the crankshaft 11 while the cam angle sensor 18 generates eight pulse signals per revolution of the cam shaft 16. In other words, provided that N is the number of signals generated by the crank angle sensor 17 while the crankshaft 11 makes a single revolution, it is designed that 2N counts of signals are generated by the cam angle sensor 18 for each revolution of the cam shaft 16. Since these crankshaft 11 and cam shaft 16 are maintained at the rotational angular velocity ratio of 2:1 by means of the transmission mechanism, when it is seen from the control mechanism 70, respective signals output from these angle sensors 17 and 18 are matched angularly 1:1 to each other.
The control unit 70, therefore, measures the rotational phase difference .theta. between the crankshaft 11 and cam shaft 16 depending on each phase difference of signals output from these angle sensors 17 and 18 and determines the necessary amount of actuation of the rotational phase adjusting mechanism 20 via the linear solenoid 56. It should be noted that the signal number N is set as a value which satisfies the following condition where the maximum actuating amount (maximum adjustable angle) of the rotational phase adjusting mechanism 20 is .theta. max..degree.CA (crank angle): N &lt;360.degree. CA/.theta.max..degree.CA.
FIG. 8 shows a structural cross section of the rotational phase adjusting mechanism 20 and outline of the rotational phase adjusting mechanism 20 is explained referring also to FIG. 8.
The rotational phase adjusting mechanism 20 is built in a housing 22 which is fixed to a cylinder head 21 of an internal combustion engine.
Approximately cylindrical cam shaft sleeve 23 is mounted and fixed with a pin 24 and a bolt 25 on the end of cam shaft 16 extending from the right hand side in the figure. In addition, at the section where the sleeve 23 supports the cam shaft 16, the sprocket 13 is engaged. The sprocket 13 is prohibited form moving in the axial direction but it can slide in the rotating or circumferential direction.
On the other hand, an approximately cylindrical sprocket sleeve 26 is fixed with a pin 27 and a bolt 28 on the sprocket 13 and further an end plate 29 is fixed to the other end of the sleeve 26.
As described above, the cam shaft sleeve 23 and the cam shaft 16 as well as the sprocket sleeve 26 and the sprocket 13 are integrated to each other and each of these is rotatable in a ring plate 31 fixed with a knock pin 30 to the housing 22.
In addition, an external toothed helical spline 32a is formed at a part on a peripheral surface of the cam shaft sleeve 23 while an internal toothed helical spline 33a is formed at a part of an internal peripheral surface of the sprocket sleeve 26. A cylinder 34 is engaged between these sleeves 23 and 26 and the external toothed helical spline 32a and the internal tooth helical spline 33a are meshed respectively with an internal toothed helical spline 32b formed on the internal peripheral surface of the cylinder 34 and an external toothed helical spline 33b formed on the peripheral surface thereof.
According to the mesh of these helical splines, each sleeve 23 and 26 is integrated rotatably with the cylinder 34 so that the revolution of sprocket 13 can be transmitted to the cam shaft 16.
In addition, since these helical splines are thus meshed, if the cylinder 34 slides in the axial direction, a thrust occurs on the meshed section so that the cam shaft 16 is slid in the rotating direction. In other words, the relative rotational phase between the sprocket 13 and the cam shaft 16 changes according to the slide of the cylinder 34 in the axial direction.
In this mechanism, the hydraulic mechanism 50 is employed as a means to slide the cylinder 34. Further, a pair of hydraulic chambers 35 and 36 are provided in the rotational phase adjusting mechanism 20.
In this connection, on the rotational phase adjusting mechanism 20 according to FIG. 8, the hydraulic chamber 35 located at the left hand side in the figure is the hydraulic chamber for the advance angle motion and the hydraulic chamber 36 at the right hand side in the figure is the hydraulic chamber for the delay angle motion. In addition, the cylinder 34 slides to right or left in the axial direction depending on the quantity of the working oil supplied to each of these hydraulic chambers 35 and 37. It should be noted that oil seals are installed appropriately at each section in the area forming these hydraulic chambers 35 and 36.
Further, in the rotational phase adjusting mechanism 20, hydraulic routes 37, 38, 39 and 40 are disposed to connect the hydraulic chambers 35 and 36 and the hydraulic mechanism 50. The hydraulic route 37 among these is a hydraulic route communicating the oil pump 52 and the spool valve 53 and the hydraulic route 38 connects the spool valve 53 and the oil pan 51. In addition, the hydraulic route 39 is a hydraulic route connecting the spool valve 53 and the hydraulic chamber 35 and the hydraulic route 40 connects the spool valve 53 and the hydraulic chamber 36. Here, the hydraulic route 40 runs from a connecting route 40a in T-shape formed in a bolt 41 for securing the ring plate 31 to the housing 22, passes through an area 40b being encircled by the bolt 41 and the cam shaft sleeve 23, reaches a hydraulic route 40c formed in the sleeve 23 and further connected therefrom to the hydraulic pressure chamber 36.
FIG. 9 illustrates in (a) through (c) thereof the structure as well as ways of movement of the spool valve 53 and the movement of spool valve 53 corresponding to the duty value control by the control unit 70 is explained referring to FIG. 9.
As depicted by FIG. 9, the spool valve 53 comprises a cylinder 54, a spool 55 sliding axially in the cylinder 54 and a spring 57 which pushes the spool 55 in the direction opposite to the driving direction by the linear solenoid 56.
In addition, a working oil supply port 58 communicating with the hydraulic route 37 (oil pump 52) of rotational phase adjusting mechanism 20 and a working oil discharge port 59 communicating with the hydraulic route 38 (oil pan 51) of the adjusting mechanism 20 are formed on the cylinder 54. Further, also in the other direction of the cylinder 54, hydraulic ports 60 and 61 communicating respectively with the above mentioned hydraulic route 39 (hydraulic chamber 35) and the hydraulic route 40 (hydraulic chamber 36) of the rotation phase adjusting mechanism 20 are formed respectively. In addition, in the advance angle movement hydraulic chamber 35 and delay angle movement hydraulic chamber 36, quantity of working oil is increased or decreased depending on a continuous change in the extent of opening of each port along with the sliding of the spool 55.
Examples of typical conditions of the spool 55 are shown in FIGS. 9A-9C and the movement of the spool valve 53 is explained below.
FIG. 9A shows an example of condition of the spool valve 53 when a current having about the duty value of "100%" value is applied to the linear solenoid 56. In this occasion, the spool 55 is driven to the right end of the cylinder 54 by the linear solenoid 56 on the spool valve 53 so that the section between the working oil supply port 58 and the hydraulic port 60 and the section between the hydraulic port 61 and the working oil discharge port 59 are connected. Under this condition, working oil is supplied through the hydraulic route 37 and 39 to the advance angle movement hydraulic chamber 35 while working oil is discharged from the delay angle movement hydraulic chamber 36 through the hydraulic routes 40 and 38. Therefore, on the rotational phase adjusting mechanism 20 according to FIG. 8, the cylinder 34 is driven toward right hand side in FIG. 8 so that the phase of cam shaft 16 is advanced relative to the sprocket 13 to establish a so-called advance angle control condition.
FIG. 9B shows in (b) an example of condition of the spool valve 53 when a current having the duty ratio of approximately "50%" is applied to the linear solenoid 56 by the control unit 70. In this occasion, in the spool valve 53, the driving force of the linear solenoid 56 and the thrusting force of the spring 57 are balanced and both the hydraulic ports 60 and 61 are closed by the spool 55. In addition, in such condition, also on the rotational phase adjusting mechanism according to FIG. 8, the balance of working oil quantity is maintained theoretically at the advance angle movement hydraulic chamber 35 and the delay angle movement hydraulic chamber 36 and the phase between the sprocket 13 and the cam shaft 16 is maintained.
FIG. 9C shows in (c) an example of condition of spool valve 53 when a current having the duty ratio of approximately "10%" is applied to the linear solenoid 56 by the control unit 70. In this occasion, in the spool valve 53, the spool 55 is pushed to the left end of the cylinder 54 by the spring 57 and the section between the working oil supply port 58 and the hydraulic port 61 and the section between the above hydraulic port 60 and the discharge port 59 are communicated respectively. In addition, in such condition, working oil is supplied to the delay angle movement hydraulic chamber 36 though the hydraulic routes 37 and 40 while working oil is discharged from the advance angle movement hydraulic chamber 35 through the hydraulic routes 39 and 38. Consequently, in the rotational phase adjusting mechanism 20 according to FIG. 8, the cylinder 34 is driven to the left hand side in FIG. 8 so that the phase of cam shaft 16 is delayed relative to the sprocket 13 to establish a so-called delay or retard angle control condition.
It is to be understood that FIG. 9 shows only typical conditions of the spool valve 53 and, actually, it changes continuously from the condition of FIGS. 9A-9C depending on the duty value of the electric current applied to the linear solenoid 56.
As explained above, in the conventional valve timing control mechanism, the working oil quantity supplied to the respective hydraulic chambers 35 and 36 for the advance angle movement and the delay angle movement can be adjusted continuously through the spool valve 53 and, in the end, the valve timing of the internal combustion engine also can be controlled to desired angles.
Here, as explained above with reference to FIG. 9B, when a current of approximately "50%" duty ratio is applied to the linear solenoid 56, both the hydraulic ports 60 and 61 are surely closed on theory and the balance of respective working oil quantity is maintained on the respective hydraulic chambers 35 and 36 for the advance angle movement and the delay angle movement.
However, actually, leakage of working oil occurs between the spool valve 53 and the hydraulic routes 39 and 40. In other words, although the phase between the sprocket 13 and the cam shaft 16 is maintained by the balancing of the working oil quantity, it is unlikely that the working oil quantity supplied to the hydraulic chambers 35 and 36 through the spool valve 53 is always reduced to 0. Even if the phase is maintained between the sprocket 13 and cam shaft 16, it means only that the working oil quantity leaking from the hydraulic route 39 or 40 and that is supplied to the hydraulic route 39 or 40 are balanced by chance.
FIG. 10 shows the relation of actuating amount (drive amount) of the linear solenoid 56 and the working oil quantity supplied to the rotational phase adjusting mechanism 20 through the spool valve 53. Regions (a) through c) of FIG. 10 correspond to respective conditions of the spool valve 53 according to FIGS. 9A-9C.
The actuating amount of the linear solenoid 56 which makes 0 the working oil quantity supplied to the hydraulic routes 39 and 40, which means the actuating amount (called the retaining output hereinafter) which balances the working oil quantity leaking from the hydraulic route 39 or 40 and that is supplied to the hydraulic route 39 or 40, corresponds to the point BP in FIG. 10. Then, at this occasion, the valve timing is retained at a constant angle.
However, the working oil quantity leaking from the hydraulic route 39 or 40 varies depending on change of viscosity according to the oil temperature, change in hydraulic pressure and so forth. Further, the working oil quantity supplied to the hydraulic route 39 or 40 also changes. Therefore, the point BP corresponding to the retaining output also is not necessarily fixed at one point; rather it always changes according to the operation conditions of the relevant internal combustion engine 10. In addition, like the conventional mechanism, if the valve timing control is executed without recognizing such change in the retaining output, not only does it becomes impossible to retain the valve timing at a constant level but also the convergent accuracy of the control will become worse.